Field of the Invention
The present invention relates generally to a wind energy extraction mechanism with the following features versus prior art embodiments having the same blade swept area: up to 2.5× overall efficiency; up to 3× power density (kg/kW); over 30 dB(A) airborne noise level reduction and elimination of infrasonic noise level generation; over 16× the reliability, due to elimination of prior art failure causal mechanisms; greatly enhanced ease of maintenance and repair; greatly simplified manufacturing, shipping, installation and erection capability; up to 3× reduction in the cost to manufacture; design robustness re dynamic wind gust, cyclic loading and sustained high-wind induced stresses on the tower Overturning Moment and Base Shear Force capabilities; up to 2.4× speed of response (yaw rate) to changing wind direction; 94% reduction of flicker; and elimination of bat and bird kill potential.
Description of Related Art
Horizontal-axis wind turbines (HAWTS) are susceptible to the Betz Limit criteria (i.e., 16/27ths), whereby they lose at least 41% of the theoretical extractable energy from wind velocity. Thereafter, the energy extraction process is solely dependent on the turbine overall efficiency. The turbine overall efficiency (ηo) consists of blade aerodynamic efficiency (ηb) times the associated mechanical efficiency (ηm) times the electrical conversion process efficiency (ηe) to produce the resultant electrical power. These efficiency terms are combined into an expression to determine the maximum extractable energy in Watts/m2 vs. the wind or current velocity. This relationship can be represented by the following expression:Watts/m2=0.50×(ρ,kg/m3×(wind vel.,m/sec.)3×Betz Limit×ηo                 where ρ=1.225 kg/m3 at sea level elevation and 68° F.        or Watts/m2=0.363×(wind vel., m/sec.)3×(ηb×ηm×ηe).        
Current wind turbine industry practice is to measure the output power from their generators without consideration of the power conditioning and conversion processes necessary for establishing grid compatibility. The reported total Watts generated is then simply divided by the rotor swept area to determine the specific energy at that wind velocity. These curves are then used in sales brochures to present documented performance capabilities. Unfortunately, this practice assumes that energy is being uniformly extracted over the entire swept area. This is not the case, as the rotor delivery torque times the rotor RPM is proportional to the input power supplied to the gearbox. The torque is composed of the summation of lift and drag forces acting at varying distances along the blade from the rotor hub to the tips. These forces are proportional to the rotational velocity2 at any particular distance from the hub. Integration of the resultant torque as a function of incremental distances along the blade will show that ˜90% of the energy extracted is being provided by the outer 30% of the rotor disc (or ˜49% of the area). This leads to the surprising conclusion that the past practice of using the entire swept area of the rotor disc to estimate the energy extracted must be reduced by half, revealing that reports of blade performance aerodynamic efficiency are ˜2× higher than is actually the case.
It is evident that wind velocities remain essentially unchanged as they pass through the inner 70% of the rotor disc of conventional wind turbines, causing large flow-field discontinuities downstream. Mixing of the highly disturbed outer flow field with that of the essentially undisturbed inner flow field generates swirling eddies downstream of the rotor.
The lift/drag ratio (CL/CD) of the blade determines its aerodynamic efficiency (ηb). This ratio is usually low, because a sufficiently strong blade cannot be created to resist the induced bending without a large section modulus. A large section modulus requires thick blade sections, typically 25% to 35% of the chord dimension, generating excessive drag. The resultant CL/CD is typically below 44, yielding an aerodynamic efficiency of 42% to 48%. A high efficiency thin section blade, such as the NACA 6412, with a CL/CD of >110, cannot be used in current large two and three blade wind turbines because of this strength requirement.
The mechanical efficiency (ηm) is primarily reflected in the turbine gearbox used to convert the 16-25 RPM of the multi-bladed rotors to 1200 RPM and higher, in order to drive one-to-four generator assemblies. These high-ratio, multistage gearboxes are required to achieve the desired 50:1 to 75:1 speed increases. As each stage is only 98.5%±0.5% efficient, a four-stage gearbox would therefore have a maximum efficiency of 92% to 96%.
The electrical efficiency (ηe) consists of both the generator efficiency and the efficiency of the associated conversion process needed to achieve the high voltage, 3-phase, 60 Hz power for grid compatibility. Typical high performance generator efficiency is 88% to 92% for either AC or DC embodiments. With a transformer, for use with an AC generator, the efficiency is typically 96.5% to 98.5%, yielding a net overall average of 88%. With use of a DC generator, with an efficiency of 88% to 92% and a solid state inverter with efficiency of 97% to 98%, the net overall average remains at 88%.
In summary: a blade efficiency of 45%, a gearbox efficiency of 96%, and a power generation and conversion efficiency of 88% yields a net system overall efficiency of 38%, or (ηb)(ηm)(ηe)=ηo. A tabulation of the performance for these prior art designs would confirm this value for the net overall efficiency and show that, once the Betz Limit is included, the total specific energy extracted is approximately 22.5% of the theoretical wind energy.
Analysis of Related Art
Existing prior art HAWT designs, such as the Vestas V80-2.0 MW wind turbine, have an overall weight of ˜1080 tons, including a rotor at 90 tons, nacelle weight with rotor of ˜150 tons, 80-meter tall tower of 170 tons, and a foundation of 760 tons. The yaw drive assemblies must be capable of handling a 150-ton nacelle load, with rotational inertia of ˜60×106 kg-m2, and are presently limited to slewing rates of ˜0.5 degree/sec.
The logistical and infrastructure required to move such large assemblies to remote wind farms demands high load capacity roadways for the transport vehicles and constitutes a major Balance of Station cost for new installations. Roadways must be engineered to support the passage of 330-ton crawler cranes and Restricted Access Vehicles (RAVs) with a very large turning radius. Also, the logistical impact with respect to traffic congestion in the site of the wind farm is severe, with up to 120 one-way trips for material and equipment per MW of installed capacity. Each tower must have a cleared 1.5-acre lay down area to permit on-site preassembly of the wind turbine rotors and placement of the blades, nacelle, and three or more tubular steel tower sections onto two separate foundations: a smaller foundation for use in rotor preassembly and the larger 760-ton foundation for the wind turbine tower itself.
HAWT wind turbines are complex structural assemblies with many eigenvalues. This complexity, coupled with little or no structural damping (<3% hysteretic), makes them highly susceptible to blade/rotor interactions with the tower structure, potentially leading to multiple modes of forced vibratory response. Near-resonance exciting forces can drive the rotor blades into large displacement amplitudes that can lead to catastrophic failure from excessive bending stresses. These vibratory amplitudes are then hard-coupled into the gearbox, and subsequently into the generator assembly. Neither of these assemblies is designed to withstand such amplified forces that, due to the lack of appreciable damping, can be multiplied by a factor of 20× or more.
HAWT rotor blades, weighing up to 30 or more tons, are extremely complex, and expensive tooling of their composite materials contributes to their high cost of fabrication. They are susceptible to catastrophic over-speeding in high wind conditions, resulting in serial failure in their redundant pitch controlled furling, blade tip air brakes, and/or main shaft braking systems. Additional catastrophic, life-threatening, failures occur when stress fractures result in to thrown blades, generating massive imbalance, leading to destruction of the component elements within the nacelle, and ultimately to blade impact with the steel tower and its resultant destruction.
The power takeoff point from the rotor main shaft is very difficult to access for performing maintenance and repair operations. The major mechanical and electrical components, including the gearbox (˜36 tons), rotor assembly (˜90 tons) and generator assembly (3 to 6 tons), are typically packed into a cramped nacelle, located 65 to 125 meters above ground. Major repairs require a 330-ton crawler crane to remove the rotor and nacelle from the tower.
HAWTs typically require heavy, multi-stage gearboxes at speed increase ratios from approximately 65:1 up to >85:1 for driving the generator(s). Both the gearboxes and the generators are highly susceptible to expensive and time-consuming failures, typically occurring within the first 2-to-3 years. The inability of the Industry to achieve theoretical lifetime goals of 20 years or more for gearbox reliability is forcing a number of wind turbine manufacturers to look at alternative approaches, such as direct-coupled low speed permanent magnet generator configurations. Unrealized goals for mean time between failures, mandating warranty periods limited to 1-2 years, and high operating and maintenance costs for gearboxes, are directly traceable to gear teeth or bearing failures caused by unexpected overload conditions, and/or failure of the lubrication system.
Existing HAWT gearbox designs are manufactured to the highest precision levels (AGMA class 12 and 13), requiring expensive tooling and time-consuming manufacturing processes to meet the design tolerances. Before shipment, a mandatory 24-hour “run-in” is performed to observe the increased particle count generated over time in the recirculated lubrication oil, to assess the efficiency of the filtration system and the degree of “wearing-in” of the gearbox itself. Although this process noticeably improves the operating efficiency of the gearbox in the relatively short time of 24 hours, once the particles are generated, they immediately initiate micro-pitting and accelerated wear.
The gearboxes are highly sensitive to loss of lubricity at temperatures above 180° F., causing the accumulation of gum and varnish, accelerating tooth wear and the buildup of backlash, and increasing failure from sudden overload conditions. The result is catastrophic tooth failure. The recirculation system must be pervasive throughout the gear train in order to mitigate hot spot generation while removing up to 360,000 BTUH from the gearbox at maximum loading. Additionally, a large 1.5 MW gearbox might typically hold 200 gallons of lubrication oil, which must be changed out semi-annually. In the event of a leak or rupture in the gearbox case, or in the associated piping recirculation and filtration system, a cleanup/remediation effort must be initiated.
These gearboxes must be sized for delivery of high levels of torque at low input speeds. They are typically sized at ˜500,000 ft-lbs with a minimum 1.25× design factor-of-safety input torque for a 1.5 MW size wind turbine rotating at a speed of ˜21 RPM. Unfortunately, this safety factor is not nearly sufficient to cope with the highly variable and very large imposed loads being transmitted into the gearbox by the rotor assemblies, which is a primary causal mechanism for inducing gearbox failures.
Gearbox failure is instigated primarily by bending or deflection of up to 1.5 meters for a 40-meter blade length, as the blades move from Top Dead Center (TDC) to Bottom Dead Center (BDC) with each rotation. Blade loading shifts rapidly as the blades attempt to accommodate a velocity profile that is spread over an elevation difference of 80 meters or more. Assuming Class 4 wind conditions and a 1/7th power wind shear exponent, with a 5.8 meter/sec wind velocity at a reference elevation of 10 meter hub height, the BDC position of an 80 meter diameter rotor on a 90 meter tall tower would be 50 meters, and its velocity would be 1.259 times 5.8, or 7.3 meter/sec. However at the TDC position, the velocity would be 1.369 times 5.8, or 7.94 meters/sec. As the theoretical energy of the wind is proportional to the velocity cubed, the watts/m2 to be absorbed is 1.287× higher at TDC. This higher force component bends the blade backwards toward the tower. As the blade circles to BDC, the blade is unloaded and bends away from the tower. This constant bending fluctuation leads to a very large number of cumulative fatigue cycles in a very short time. Typically, accumulated fatigue cycles over a period of one year would exceed ten million, assuming a nominal 21 RPM rotating speed for an 80 meter diameter rotor with a Tip Speed Ratio of 6.4, and an average annual wind velocity of 7.3 meters/sec. The magnitude of the fatigue cycle is equivalent to a 1.5 MW size wind turbine assembly operating with a ±28.7% “torque ripple” pulse per blade at a frequency of approximately one Hertz. This torque ripple alone can induce early gearbox failure. However, coupled with wind gusts of a similar magnitude (or ±28.7% of average wind speed) the cumulative effect of the second term would result in doubling the velocity. This yields an 8× increase in the fluctuating wind energy (due to the velocity cubed effect) or 2.30 times the nominal design loading with each cycle. Providing a sound mechanical design for this overload condition is a formidable challenge, and appears to be one of the most likely reasons that a large number of wind turbines are laying idle, awaiting repair.
At the 6.0× to 6.4× tip speed ratios of current turbines, the resultant wind velocity is a primary causal mechanism of noise generation and generation of violent turbulent eddies and swirl effects off the blade tips. The separation of the trailing edges of the blade generate mid-frequency audible tones—the “swoosh, swoosh” noise of the blade passing in front of the tower. Because acoustic noise generation increases as the fifth power of RPM, a doubling of RPM yields a 32-dBA increase. Blade tip speeds are proportional to the number of blades; a 3-bladed wind turbine with a TSR of 6.4 and an 80 m rotor diameter yields an equivalent RPM of 21 and a tip speed of 200 mph. Obviously, the bird kill potential for these prior art designs is also quite high.
The turbulent eddy and swirl of these prior art HAWT designs exacerbate both the downwind and crosswind effects of adjacent wind turbine assemblies, and induce higher levels of discontinuity to the incident wind of the partially shaded adjacent downwind and crosswind turbines. Current practice is to space these adjacent wind turbines at a distance of 10× rotor diameters for downwind turbines and 5× rotor diameters for crosswind turbines. It is a well-known phenomenon that both downwind and crosswind turbines are less reliable than the front row lead turbines in a large wind farm.
The low reliability of gears, blades, bearings and generators makes it difficult to offer more than a one-year warranty, generates large warranty payments caused by failed components, and creates substantial downtime and inability to meet mean-time-to-repair (MTTR) expectations. The failure rate data compiled by the European Wind Energy Association (EWEA) shows an average of under 7400 hours mean-time-between failure (MTBF) occurring in a sample population of 6000+ wind turbines, with an average MTTR of 17 days or more. This indicates that, during a period of one year, any individual wind turbine can be expected to be shut down for repairs for an average period of up to 17 days.
Many systems cannot operate in a cost effective manner in less than US DOE/National Renewable Energy Laboratory Class 4 wind conditions (˜5.8 m/sec ref 10 m elevation). Additionally, purchases of new wind turbines declines rapidly without the availability of subsidized support from the federally-mandated Production Tax Credit (˜1.9¢/kWh). Finally, profitable operation is not generally possible, even in Class 4 winds, without Investment Tax Credits and allowances for Double Declining Depreciation and amortization schedules of up to 30 years.
Existing prior art wind turbines suffer from a number of liabilities other than their exceptionally poor efficiency, poor economics, and serious reliability problems. These additional liabilities are related to sensory impact on the local population residing in the vicinity of the wind farm. They include both audible noise level generation and inaudible infrasonic noise, which can travel for many miles; unsightliness and obstruction of view from the towers, rotor blade flicker; and the kill-rate of birds and bats. Setback requirements in populated areas are typically a minimum of five to ten rotor diameters, using large tracts of land. For these reasons, a prevailing “Not In My Back Yard” (NIMBY) set of objections make the permitting processes for gaining wind farm site acceptance a generally long and arduous affair.